Broad use of natural gas and hydrogen as fuels and energy carriers will provide the US with better energy security, return major economic, environmental, and health benefits, and will minimize the climate change impact of greenhouse gas emissions. Natural gas and hydrogen couple into any realistic model of the “renewable and sustainable natural gas-hydrogen-electricity economy” in an integrated and critical manner. For example, because of their relatively higher volumetric energy density and gravimetric energy density, liquid natural gas (LNG) and liquid hydrogen (LH2) are superior to compressed natural gas or hydrogen for cost-effective storage and delivery means to end users. Further, for aircraft, heavy-duty vehicles such as class 7/8 trucks, marine vessels, and train engines, LNG or LH2 are superior on-board fuel storage options compared to compressed gas storage techniques.
The ideal specific power input for a liquefier of a certain rate of cryogen production depends on the gas to be cooled and liquefied. The ratio of the ideal minimum work input per unit mass of gas to the real work input per unit mass of gas for a practical liquefier is called the figure of merit (FOM). Currently the majority of LH2 is commercially liquefied via the nitrogen pre-cooled Claude cycle while LNG is primarily commercially liquefied by turbo-Brayton, cascade, or mixed refrigerant cycles. Existing large-scale conventional liquefiers for either cryogen can only achieve FOMs of ˜0.35 and the FOMs decrease steadily for smaller-scale liquefiers. The increased price of these fuels due to low FOM presents a major technical barrier to the adoption of either LNG or LH2.
There is a great need for more efficient and less expensive cryogen liquefaction methods. At smaller liquefier plants such as those needed to supply distributed refueling stations, this technology gap must be filled to enhance adoption of LNG and LH2 into the US$675 billion/year transportation fuels market. Magnetic refrigeration based on the magnetocaloric effect was identified as a potential method for more efficient refrigeration because it uses solids rather than gases as refrigerants and magnetic field changes instead of gas pressure changes to create a thermodynamic cycle.
An early advancement toward efficient magnetic liquefiers was first disclosed by Barclay and Steyert (1982) in U.S. Pat. No. 4,332,135 wherein certain magnetic materials were simultaneously used as active refrigerants and thermal regenerators as an active magnetic regenerator (AMR). An AMR can be made to execute an active magnetic regenerative refrigeration cycle comprised of four steps: demagnetization of the AMR from high to low field with no heat transfer fluid flow; a hot-to-cold flow of heat transfer fluid at constant low field; magnetization of the AMR from low to high field with no fluid flow; and a cold-to-hot flow of heat transfer fluid at high field. After step 2 the cold heat transfer fluid absorbs heat from a thermal load before executing the cold to hot flow in step 4. The hot heat transfer fluid rejects heat to the heat sink after step 4 before it executes step 2. The required thermodynamic work for the AMR cycle is provided via differences in magnetic forces between magnetic materials in the regenerators and the high magnetic field as the materials enter and leave the field. Detailed analysis shows that good AMRL designs can achieve FOMs of 0.5-0.7 largely by eliminating large inefficiency of gas refrigerant compression typical of conventional hydrogen or natural gas liquefiers.
Currently, reciprocating dual regenerator AMRL designs dominate existing prototypes where piston-shaped layered magnetic regenerators are reciprocatively moved into and out of high magnetic field regions with reciprocating, intermittent heat transfer fluid flows phased to execute the heat transfer steps of the AMR cycle during flows axially through each porous regenerator. Reciprocating AMRL designs have inherent features that increase operational complexity. For example, each regenerator only provides cooling to an external thermal load or process stream load during heat transfer fluid flow from the hot end to the cold end of the demagnetized regenerator in one of four steps of an AMR cycle. To provide continuous cooling of a process stream in a liquefier, four different regenerators with movement phased 90° apart in and out of two solenoidal magnets must be integrated into an AMRL to sequentially provide continuous cooling of the process stream or other thermal loads. Further, the large magnetic forces between the magnetic regenerators and the magnetic field in reciprocating designs are difficult to react even in a dual regenerator configuration. Resultant heat generation in the persistent-mode magnets can overload the cryocooler used to keep magnets superconducting at about 4 K for NbTi, a widely used superconducting wire. The intermittent reciprocating heat transfer fluid flows and axial drive motions in reciprocating designs are increasingly difficult to implement as operating frequencies increase from 0.1-0.25 Hz to 1-2 Hz which is desirable to linearly increase cooling power per kg of refrigerant.
The complicating design features of reciprocating AMRLs can be eliminated by using a rotary regenerative wheel configuration. Wheel designs have a continuous rim comprised of one or more layers of porous ferromagnetic refrigerants with cascading Curie temperatures and means for continuously radially or axially flowing heat transfer gas only through sections of the rotating porous wheel rim in the high and low magnetic field regions and within a stationary hermetic housing. The simultaneous execution of all four steps of the AMR cycle at different sections of a rotating magnetic wheel within the stationary housing and fixed high and low magnetic fields naturally gives several significant AMRL design advantages.
Some advantageous features of rotary magnetic wheel AMRL designs were originally described by Barclay (1983) in U.S. Pat. No. 4,408,463A. Differentiating features include: i) steady-state temperatures varying from the heat rejection temperature THOT to lowest cooling temperature TCOLD occur at various locations in the AMRL; ii) continuous cooling is provided to external thermal loads and process loads in heat exchangers with continuous heat rejection in the hot exchanger; iii) steady pressures exist at any point in the heat transfer flow circuits with the highest pressure at discharge of the continuously circulating heat transfer fluid pump and lowest at the suction side of the pump; iv) attractive magnetic forces on segments of the magnetic wheel rotating out of the high magnetic field region are continuously reacted by slightly smaller attractive magnetic forces on segments of the magnetic wheel rotating into the high magnetic field region which gives AMR cycle thermodynamic work as a constant torque provided to rotate the magnetic wheel by a continuous drive mechanism; and v) the persistent mode superconducting magnets have fixed magnetic flux from the rotating magnetic material so the cryocooler thermal load to cool the magnets remains constant during the AMR cycle. Such features make continuous wheel type AMRL designs highly attractive for cryogen liquefaction applications for any temperature span from 4 K to 280 K.
However, the major challenge faced by the rotary wheel design is restricting primary heat transfer fluid flows between the rotating porous magnetic wheel and the stationary hermetic housing that contains the wheel. In short, a properly sealed rotary wheel AMRL device allows radial or axial heat transfer fluid flows through sections of rotating porous magnetic wheel within the high and low fields regions, while restricting undesirable leakage flow over, under, around, or circumferentially along the rotating bed in the small space between the rotating wheel and the stationary housing.
Solutions to restricting fluid flow between rotating surfaces have been in existence since the early 1900s. Applications include shaft seals for steam turbines and surface seals in many turbine-type devices such as steam and jet turbines. Early solutions were limited by material availability and manufacturing techniques. The need for more efficient aircraft turbine engines drove rotating seal advancement in the early 1960s. In U.S. Pat. No. 2,963,307, Bobo (1960) describes sealing axial pressure gradients using a combination of labyrinth and honeycomb surfaces to limit fluid flow across the seal even when materials respond to temperature driven eccentricities. In U.S. Pat. No. 3,231,284, Peickii et al. (1962) describe sealing pressure gradients across curved surface faces using low cost materials and manufacturing of this mechanical seal. In U.S. Pat. No. 3,339,933, Foster (1967) further improves on traditional axial-flow seal designs by incorporating a bonded high-temperature oxide material with thermal insulation characteristics onto one seal surface to reduce distortion and cracking in the seal region. Similarly, with U.S. Pat. No. 3,537,713, Matthews et al. (1970) describe using a rub-tolerant surface to mate with rotating labyrinth seals to help maintain a tight seal while limiting friction at the mating surface.
With U.S. Pat. No. 4,218,066, Ackermann (1980) advances surface seal techniques by implementing a canted, honeycomb cell surface with an abrasion resistant coating that increases aerodynamic resistance to flow across the seal. In U.S. Pat. No. 5,110,033, Noone et al. (1992) build upon early brush seal technology to accommodate thermal and centrifugal growth or displacement with a segmented design. Sanders (1997), in U.S. Pat. No. 5,603,510, discusses a specific variable clearance seal design that allows for seal gap clearances to adjust to changing structural dimensions. Patents U.S. Pat. Nos. 6,030,175A and 6,131,910A by Bagepalli et al. (2000) modernize previous annular seal designs by hybridizing labyrinth and brush seal techniques into one circumferential axial-flow seal for retrofitting existing labyrinth-only axial-seal systems and providing variable gap clearance needs associated with thermal and centrifugal displacement of rotating parts.
In terms of rotating face seals, or radial face seals, early attempts to seal shafts beyond the narrow annular type axial seals described above looked like Voitik's (1967) labyrinth-type face seal in U.S. Pat. No. 3,333,856. Finger seal arrangements adapted from Honeycutt, Jr. et al.'s (1987) sliding finger seal design in U.S. Pat. No. 4,645,217 and Bridges et al.'s (1987) brush seal design in U.S. Pat. No. 4,678,113 create a labyrinth type sealing surface shaped more like flat discs instead of early cylindrical designs. Ide (1995) in U.S. Pat. No. 5,385,409 develops a gap-type, non-contacting face seal that can accommodate seal face separation at startup to prevent damage to the seal faces before operational pressures are obtained. Braun at al.'s (2008) elegant hybridization of finger type axial and radial face seal in US Patent Application 2008/0122183 A1 leverages a multilayered labyrinth type approach.
Missing in the patent literature are solutions that provide means for sealing fluid flow in all three possible leakage paths within a free-space packing between a rotating torus and a stationary hermetic housing, including radial, axial, and circumferential leakage flows while allowing required primary axial or radial flows through the porous toroidal mass in certain regions of the its rotation. The present disclosure is directed, in part, to controlling and directing the fluid flow in leakage paths.